Apparatus for eliminating whirl instability in a gas supported bearing

ABSTRACT

A gas supported bearing includes a first stationary bearing element having a longitudinal axis and a first bearing surface. A second rotating bearing element is coaxially aligned with respect to the first bearing element and includes a second bearing surface. A pneumatic load ramp is formed in one of the bearing surfaces to apply an asymmetric load to either bearing element to eliminate bearing whirl instability.

This application is a Continuation-in-Part patent application of U.S.patent application Ser. No. 07/705,547, filed May 24, 1991 now abandonedwhich is a continuation of Ser. No. 568,416, filed Aug. 16, 1990, nowU.S. Pat. No 5,019,738, issued May 28, 1991.

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates to gas supported bearings, and more particularlyto high precision, high speed gas supported bearings.

2. Description of the Prior Art

FIGS 1A and 1B depict a polygon mirror scanning system having a rotatingpolygon mirror 10 mechanically coupled to a rotating cylindrical shaft12. The lower end of shaft 12 is rotatably coupled to the scannerhousing 14 by ball bearing 16; the upper end of shaft 12 is coupled tohousing 14 by ball bearing 18. Seals 20 minimize the circulation ofliquid lubricant discharged from ball bearings 16 and 18 during highspeed operations.

A permanent magnet 22 is rigidly coupled to shaft 12. When energized,motor field windings 24 interact with magnets 22 to rotate shaft 12 andpolygon mirror 10.

Such prior art ball bearing supported motor driven loads respond todimensional irregularities in the ball and race assemblies of the ballbearing and adverse interaction with the liquid lubricant can generatepolygon mirror scanning errors of ten arc seconds or greater dependingon the spacing between the two supporting bearing assemblies. Even whenselected elements of the ball bearing scanning assembly are custommachined and custom fitted, scanning errors generally cannot be reducedbelow about five arc seconds. Lube redistribution can contribute torotational period instability (velocity stability).

The unpreventable circulation of liquid lubricant discharged by the ballbearings enters the interior of housing 14, contaminates the reflectivefacets of polygon mirror 10, particularly along the leading edge of eachfacet, and thereby degrades the reflectivity of the mirror.Periodically, the individual facets of polygon mirror 10 must be cleanedto remove contaminating lubrication.

The prior art herringbone bearing assembly illustrated in FIG. 2includes a cylindrical bore 26 and a shaft 28. Shaft 28 includesdiscrete herringbone patterns designated by reference numbers 30 and 32.Each herringbone pattern must be formed with the highest possibleprecision in the outer surface of shaft 38. As illustrated by the edgeof the sectional view of shaft 28 as designated by reference number 34,approximately fifty percent of the shaft surface area within theherringbone pattern area is removed so that only approximately fiftypercent of the remaining shaft surface can form a load supportingsurface between the rotating shaft and the uninterrupted, cylindricalsurface of the sleeve bore 26. This sharply limited load supportingsurface area drastically reduces the load supporting forces or bearingstiffness generated between shaft 28 and sleeve 26. As a direct result,the closely spaced surfaces of shaft 28 and sleeve bore 26 do not liftoff and become airborne until the grooves become pressurized. From 0 RPMto lift off velocity, these two surfaces operate as a contact bearingand mechanically rub against each other generating significantfrictional forces and bearing surface wear.

The herringbone air bearing depicted in FIG. 2 relies upon the airpumping action generated by the interaction between the relativelyrotating sleeve bore 26 and shaft 28. Such pumping forces generate aflow of pressurized air in the direction indicated by arrows 36 flowingupward through the bearing surface and are discharged from air dischargeport 38. Once appropriate pressurization has been established by therotating sleeve assembly, the sleeve bore 26 becomes airborne relativeto the crowned top of the shaft 28. Until liftoff occurs, the top ofshaft 28 rubs upon and can create surface wear at the interface betweenthe top of shaft 28 and the base of air discharge port 38.

Another disadvantage of herringbone air bearings of the type depicted inFIG. 2 is that they must be operated in a vertical orientation.Deviation from the desired vertical alignment on the order of tendegrees of inclination can create rapid bearing surface wear and canresult in failure of the herringbone bearing assembly.

The high level of mechanical precision required to create theherringbone pattern in the surface of bearing shaft 28 contributes to ahigh manufacturing cost.

SUMMARY OF THE INVENTION

It is therefore a primary object of the present invention to provide ahigh precision, gas supported bearing to rapidly and at low velocitybuild up high level bearing stiffness on the order of about 30,000 to50,000 pounds per inch capable of operation over a wide speed range.

Another object of the present invention is to provide a high precision,gas supported bearing which produced lift off of the bearing sleeverelative to the bearing shaft at relatively low velocity.

Another object of the present invention is to provide a high precision,gas supported bearing which generates extremely high level bearingstiffness forces enabling operation of the bearing horizontally,vertically or in any intermediate attitude over a broad speed range.

Another object of the present invention is to provide a high precision,gas supported bearing which is capable of bidirectional operation.

Another object of the present invention is to provide a high precision,gas supported bearing capable of operation as a closed system to reduceproblems caused by pumping debris into the bearing air gap.

Another object of the present invention is to provide a high precision,gas supported bearing which is capable of a minimum of 20,000 start/stopcycles.

Another object of the present invention is to provide a gas-supportedbearing having a pneumatic load ramp formed in one of the bearingsurfaces to apply an asymmetric load to the bearing at a location fixedrelative to either bearing element to eliminate bearing whileinstability.

Another object of the present invention is to provide a gas supportedbearing where the asymmetric load generated by a pneumatic load rampincreases with increasing bearing operating RPM to eliminate bearingwhirl instability at various bearing operating RPM's.

Briefly stated, and in accord with one embodiment of the invention, agas supported bearing includes a cylindrical bearing sleeve having alongitudinal axis and a cylindrical inner surface including a firstbearing surface. The bearing also includes a cylindrical shaftpositioned coaxially within the bearing sleeve. The bearing shaftincludes a cylindrical outer surface having a second bearing surface. Arelative rotational velocity is established between the bearing sleeveand the bearing shaft to generate a bearing supporting force along abearing overlap zone where the first bearing surface overlaps the secondbearing surface. The bearing overlap zone includes a defined lengthalong the longitudinal axis. A pneumatic load ramp is formed in one ofthe bearing surfaces and includes a length along the longitudinal axisless than the length of the bearing overlap zone for applying anasymmetric load to the bearing at a location fixed relative to eitherthe sleeve or the shaft to eliminate bearing whirl instability.

DESCRIPTION OF THE DRAWINGS

The invention is pointed out with particularity in the appended claims.However, other objects and advantages together with the operation of theinvention may be better understood by reference to the followingdetailed description taken in connection with the followingillustrations, wherein:

FIG. 1A is a sectional view of a prior art polygon mirror scannerincluding a rotating cylindrical shaft supported at each end byconventional ball bearings.

FIG. 1B represents a partially cutaway, perspective view of the scannerdepicted in FIG. 1A.

FIG. 2 represents a sectional view of a prior art herringbone gassupported bearing.

FIG. 3 represents a sectional view of a polygon mirror scanner includinga self-pressurizing gas supported bearing of the present invention. Inthis embodiment of the invention, the cylindrical bearing shaft isrotated relative to a fixed cylindrical bearing sleeve.

FIG. 4A represents a sectional view of a polygon mirror scannerincluding a self-pressurizing gas supported bearing of the presentinvention. In this embodiment of the invention, the cylindrical bearingsleeve rotates about a stationary cylindrical shaft.

FIG. 4B is a partially cutaway top perspective view of the polygonmirror scanner depicted in FIG. 4A.

FIG. 5A represents a partially cutaway, illustrative sectional viewshowing the smooth bearing surfaces typical of prior art bearingstructures.

FIG. 5B represents a partially cutaway, illustrative sectional viewshowing the bearing surfaces of the present invention including adefined surface roughness.

FIGS. 6A, 6B and 6C represents a series of views used to definemechanical engineering terms including Roughness Average R_(a).

FIG. 7 illustrates a cylindrical bearing shaft and sleeve where thegeometry of the bearing gap changes as a result of taper error.

FIG. 8 illustrates a cylindrical bearing shaft and sleeve including bellmouth geometric errors at each end of the bearing assembly.

FIG. 9 illustrates the shaft and sleeve of a cylindrical bearingincluding a non-uniform air gap caused by a bowed shaft.

FIG. 10 illustrates a bearing assembly including a single shaft with twospaced apart sleeve elements where the gap between the shaft and sleevefor each bearing element is non-uniform.

FIG. 11 illustrates the shaft and sleeve of an air bearing illustratinggeometric errors due to barrel effect.

FIG. 12 illustrates the shaft and sleeve of an air bearing illustratinggeometric errors in the central portion of the sleeve bore.

FIG. 13 is a sectional view of a cylindrical air bearing assemblyincluding a shaft and sleeve illustrating the variation in gap dimensioncaused by geometric errors.

FIG. 14 defines the bearing aspect ratio as determined by the ratio ofbearing shaft diameter to bearing sleeve length.

FIG. 15A illustrates the relative area of lands and grooves of a ceramicshaft.

FIG. 15B is a partially cutaway sectional view illustrating the landsand grooves of a ceramic shaft.

FIG. 16A illustrates a random cross-hatched roughness pattern added tothe surface of a metal bearing element.

FIG. 16B is partially cutaway sectional view of the bearing elementillustrated in FIG. 16A, illustrating alternating lands and grooves.

FIG. 17A illustrates random, circumferential grooves in the surface of ametal bearing shaft.

FIG. 17B is a partially cutaway sectional view of the bearing elementdepicted in FIG. 17A, showing the relationship between bearing lands andgrooves.

FIG. 18 illustrates an embodiment of the invention including a pair ofcounter-rotating sleeves on a non-rotating bearing shaft.

FIG. 19 illustrates a sectional view of an embodiment of the inventionincluding spherical bearing surfaces.

FIG. 20 illustrates a sectional view of an embodiment of the inventionincluding bearing surfaces formed as conic sections.

FIGS. 21A, B and C represent sectional views of self-pressurizing gassupported bearings configured as axial thrust control bearings.

FIG. 22 is a perspective view illustrating a pneumatic load ramp formedin the shaft of a self-pressurizing gas supported bearing.

FIG. 23 is a sectional view of the bearing shaft illustrated in FIG. 22,taken along section lines 23--23.

FIG. 24 is a partially cutaway perspective view of a stationary bearingsleeve incorporating a pneumatic load ramp of the present invention.

FIG. 25 is a sectional view of the bearing sleeve illustrated in FIG.24, taken along section lines 25--25.

FIG. 26 is a ninety degree sectional view of a self-pressurizing gassupported bearing including the pneumatic load ramp of the presentinvention formed in the surface of a bearing shaft.

FIG. 27 shows the pneumatic load ramp of the present invention formed asthree separate, spaced apart ramp elements.

DESCRIPTION OF THE PREFERRED EMBODIMENT

In order to better illustrate the advantages of the invention and itscontributions to the art, a preferred hardware embodiment of theinvention will now be described in some detail.

Referring now to FIG. 3, a self-pressurized gas supported bearingincludes a cylindrical bearing sleeve 40 having a longitudinal axis 42and a cylindrical inner surface 44 which forms a first bearing surface.

A cylindrical bearing shaft 46 is positioned coaxially within bearingsleeve 40 and includes a cylindrical outer surface 48 which forms asecond bearing surface. The upper end of shaft 46 is rigidly coupled tosupport a load such as a rotatable polygon scanning mirror 50.

The lower end of shaft 46 is coupled to an annular magnet assembly 52which forms a part of an axial thrust bearing assembly which alsoincludes a non-rotating annular magnets 54 and 56. As shown in FIG. 3,magnets 52, 54 and 56 are positioned with opposing poles to createmagnetic repulsion forces both above and below rotating magnet 52. Theseessentially equal magnetic repulsion forces maintain an essentiallyfixed spacing on the order of about 0.030 inches between magnets 52 and54 as well as between magnet 52 and magnet 56 to maintain an essentiallyfixed axially alignment for both shaft 56 and a load such as polygonmirror 50.

Although this particular axial thrust bearing design works well,numerous other types of axial thrust bearing assemblies well known toone of ordinary skill in the art could easily be substituted for themagnetic axial thrust assembly illustrated in FIG. 3.

Drive means in the form of an electric motor 58 is coupled to establisha desired relative rotational velocity between bearing sleeve 40 andbearing shaft 46. In FIG. 3, electric motor 58 includes a permanentmagnet assembly 60 which is rigidly coupled to the outer surface ofshaft 46 and a field winding 62 which is rigidly coupled to thenon-rotating assembly housing 64.

Referring now to FIGS. 4A and 4B, another embodiment of theself-pressurizing gas supported bearing of the present inventionutilizing a fixed shaft and a rotating sleeve will now be described indetail.

FIG. 4A illustrates that bearing shaft 66 includes upper and lower endswhich are rigidly coupled to cylindrical apertures in opposing ends ofhousing 64. Rotating bearing sleeve 68 is positioned coaxially outsideof the bearing shaft. A load such as a rotating polygon mirror ismechanically secured to rotating sleeve 68 by a plurality of screws 70.

Axial thrust control magnet assemblies 72 and 74 are comparable to theassembly described in FIG. 3 are located at the upper and lower ends ofrotating sleeve 68 to maintain essentially fixed axial or longitudinalposition of sleeve 68 relative to shaft 66. Each magnet assemblyincludes a non-rotating magnet 76 and a rotating magnet assembly 78which is rigidly coupled to each end of rotating sleeve 68.

Referring now to FIGS. 5A and 5B, prior art self-pressurizing gassupported bearings virtually universally use extremely smooth, highlypolished bearing surfaces designated by reference number 80 in FIG. 5A.FIG. 5B illustrates the corresponding sleeve and shaft bearing surfacesof the self-pressurizing gas supported bearing of the present inventionwhich necessarily include a quantified degree of roughness to theopposing bearing surfaces as illustrated by the bearing surfacesdesignated by reference number 82.

Referring now to FIGS. 6A, 6B and 6C, specific engineering termsrelating to measurement of surface texture will now be reviewed toassist in defining the relevant roughness characteristics of the bearingsurfaces 82 of the present invention.

Surface texture is generally recognized in mechanical engineering toinclude the following four characteristics:

1. Roughness--the finer irregularities in surface texture;

2. Waviness--the more widely spaced component of surface texture;

3. Lay--the direction of the predominant surface pattern; and

4. Flaws--the unexpected, unwanted surface texture.

FIG. 6A represents a perspective view of a segment of a surface intendedto represent a reasonably linear surface. FIG. 6B represents an enlargedpartial sectional view of the surface illustrated in FIG. 6Aillustrating the surface texture characteristic of waviness. FIG. 6Crepresents an enlarged sectional view taken from FIG. 6B illustratingthe absolute variations of surface height including peaks and valleys.Roughness Average designated by the symbol R_(a) is graphicallyillustrated in FIG. 6C and is defined in mechanical engineering terms asmeaning the arithmetic average of the absolute values of measuredprofile height deviations taken within the sampling length and measuredfrom the graphical centerline. Roughness Average or R_(a) is universallyexpressed in micrometers.

The R_(a) parameters and limitations of the bearing surfaces of presentinvention will be expressed in terms of R_(a). R_(a) is further definedby MIL STD-10A dated Oct. 13, 1955 and is measured by readily availablecommercial test equipment such as Surftest Model 211 SurfaceProfilometer manufactured by Mitutoya of Japan.

The self-pressurizing gas supported bearing of the present invention canbe implemented by using a variety of materials for the sleeve and shaftelements of the present invention as illustrated in FIGS. 3, 4 and 5B.Specially, the material combinations listed in Table 1 below have foundto provide successful bearing surfaces for use in the present invention:

                  TABLE 1                                                         ______________________________________                                        SLEEVE             SHAFT                                                      ______________________________________                                        1.   Steel             Ceramic                                                2.   Hard anodized aluminum                                                                          Steel                                                  3.   Hard anodized aluminum                                                                          Ceramic                                                4.   Hard anodized aluminum                                                                          Hard anodized aluminum                                 5.   Ceramic           Ceramic                                                6.   Ceramic           Steel                                                  ______________________________________                                    

To create a high precision, self-pressurizing gas supported bearingaccording to the present invention, four different parameters must becarefully controlled within defined limits. When such parameter controlis properly implemented, a high precision bearing assembly can becreated yielding non-repeatable errors significantly less than five arcseconds and typically equal to or better than one arc second at ambienttemperature. At higher temperatures on the order of 62° C. (140° F.),such non-repeatable errors can typically be controlled to a level equalto or less than about two to three arc seconds. Such accuracy isvirtually an order of magnitude better than has been attained by ballbearing assemblies typically used in high precision scanning systemsincluding photocopy machines, laser printers and related devices.

The first parameter which must be controlled to produce the highprecision bearing of the present invention is referred to as the bearinggeometry which includes the subcategory parameters of straightness,roundness and size uniformity. With cylindrical bearing configurations,the parameter of size uniformity includes both barrel and taper errors.

Referring now to FIGS. 7-12, the various types of geometry and geometricerrors will now be briefly discussed although such terminology is wellknown to those of ordinary skill in the art.

FIG. 7 represents a relatively straight shaft 84 which is disposedwithin a sleeve 86 having a tapered bore. Such taper size uniformityvariations create a non-uniform gap between shaft 84 and sleeve 86 whichcan, if excessive, degrade the performance of the bearing of the presentinvention.

FIG. 8 illustrates a straight shaft 84 positioned within sleeve 86having a bore demonstrating bell mouth size uniformity variation at eachend.

FIG. 9 illustrates a bowed shaft 84 within sleeve 86 having a highlyaccurate bore, creating a non-uniform air gap due to the lack ofstraightness of shaft 84.

FIG. 10 illustrates a dual bearing assembly defined by a single shaft 84and a pair of sleeves 86. A non-uniform gap is created as a result ofdifferential bore diameters between left hand sleeve 86 and right handsleeve 86.

FIG. 11 illustrates a bearing assembly having a straight shaft with asleeve 86 having a barrel configuration geometric error creating anon-uniform gap between the shaft and sleeve.

FIG. 12 illustrates a bearing assembly having a non-uniform gap producedby a barrel configuration error within the central portion of the boreof sleeve 86.

In implementing the present invention, the geometry of the shaft andsleeve elements of the present invention must be controlled withindefined limits to create the desired, inventive function of the presentinvention. Specifically, as illustrated by FIG. 7, the geometryvariations within the gap of the bearing assembly of the presentinvention must be controlled to limit the gap between the sleeve and theshaft to a total distance across the bearing of approximately equal toor greater than about 100 microinches up to less than about 350microinches. As illustrated in FIG. 7, the first radial component of theoverall gap dimension designated by reference number 88 is added to thesecond radial component of the gap dimension designated by referencenumber 90. The sum of the gap dimensional contributions designated byreference numbers 88 and 90 should fall within the range of betweenabout two hundred microinches to about three hundred microinches tocreate a successful properly functioning gas bearing from the materialslisted in Table 1.

For newly discovered materials not expressly listed in Table 1, theminimum gap dimension of one hundred microinches should apply, but themaximum gap dimension could conceivably increase above the typical threehundred microinch upper limit if such previously untested material cancause the two opposing bearing surfaces to lift off and become airborneat a velocity below the bearing operating velocity. One bearing elementbecomes airborne relative to the other bearing element when sufficientbearing stiffness is created between the relatively rotating bearingsurfaces and the intervening gaseous layer moves the two surfaces out ofmechanical contact with each other.

FIG. 13 represents a sectional view of shaft 84 and sleeve 86 furtherillustrating the diametrically spaced apart gap elements 88 and 90 whichmust be controlled to fall within the limits of the present invention.

Referring now to FIG. 14, another geometry-related aspect of the presentinvention which must be controlled relates to the ratio of the shaftdiameter designated by dimension "X" to sleeve length. The aspect ratioof sleeve length to shaft diameter for materials of the type listed inTable 1 must typically substantially equal three to one and mostpreferably equal about four to one or greater. Although there is noupper limit on the maximum aspect ratio of sleeve length to shaft ratio,as a practical matter, substantially longer sleeve lengths createserious difficulties relating to maintenance of the geometric gap limitsof the present invention. As the length of the sleeve increases for agiven shaft diameter, it becomes more difficult and much more expensiveto maintain the required cylindricity necessary to implement the presentinvention.

The next parameter which must be controlled to implement theself-pressurizing gas supported bearing of the present invention is thesurface texture or R_(a) of the bearing surfaces. To achieve proper highaccuracy, low wear characteristics of the present invention, the sum ofthe R_(a) contributions from both the sleeve and the shaft must beapproximately equal to or greater than about a minimum R_(a) of eighteen(relatively smooth) and a maximum R_(a) of 60 (relatively rough).

For the materials listed in Table 1, it has been found that although thesum of the R_(a) contributions of the sleeve R_(a) plus the shaft R_(a)must approximately equal or exceed eighteen, it has also be found thatif the R_(a) of either the sleeve or the shaft falls below a minimumR_(a) rating, the bearing of the present invention will not operateproperly. For example, the minimum R_(a) for the sleeve must be aboutequal to or greater than an R_(a) of four while the minimum R_(a) of theshaft must be approximately equal to or greater than about seven. In allcases, the R_(a) total must either approximately equal or exceedeighteen. For an R_(a) rating for the sleeve on the order of about four,the R_(a) contribution of the shaft must be approximately equal to orgreater than fourteen. Similarly, for a minimum R_(a) shaft rating onthe order of about seven, the sleeve R_(a) must be about equal to orgreater than about eleven.

When the R_(a) of the sleeve plus the shaft falls below about an R_(a)of eighteen, the bearing stiffness decreases and the well knownphenomena of bearing coning increases, causing wobble of the load. Inthe rotating polygon mirror scanner embodiment illustrated in FIGS. 3and 4, such coning errors translate into increased angular deviations ofthe output beam of the optical scanner. A similar phenomena occurs whenthe overall R_(a) rating increases above about sixty or when thegeometric errors exceed about one hundred microinches total.

The last parameter which must be controlled to create theself-pressurizing gas supported bearing of the present invention relatesto the ratio of randomly distributed depressions in the bearing surfaceto the overall area of each bearing surface. As will be explained below,the bearing surface of both the sleeve and shaft must be speciallyselected and treated to include a predetermined minimum and maximumratio of depressions capable of creating air reservoirs for the overallbearing surface area. In certain materials, these air reservoir-formingdepressions take the form of grooves, cross hatching patterns orpockets.

Referring now to FIGS. 15A and 15B, the ratio of depressions or groovesto overall bearing surface for a ceramic shaft application will now beexplained in detail.

The shaded sections of FIG. 15A represent the raised areas or lands 92of an alumina ceramic shaft together with the intervening low spots orgrooves located adjacent to each land. The grooves or depressed areas 94create air reservoirs or pockets which are critical to the properfunction of the bearing of the present invention.

For the bearing surface materials designated in Table 1, the overallarea of the pockets or air reservoirs must be equal to or less thanabout fifty percent of the overall surface area of the bearing. Foroptimum performance levels, the area of the pockets or air reservoirsshould fall generally within the range of about thirty to fifty percentof the overall bearing surface area such that the area of the lands 92of the bearing representing the load bearing surface encompassesapproximately fifty to seventy percent of the overall surface area.

For ceramic bearing materials of the type schematically illustrated inFIG. 15, microphotographs enlarged to approximately 300X permit visualinspection of the air pockets and lands of the ceramic material andfacilitate computation of the ratio of the air pockets to the overallceramic surface area.

Referring now to FIG. 16, FIG. 16A illustrates a partially cutawayperspective view of the surface of a steel or aluminum shaft or sleevewhich has been treated by a honing process, or with abrasive technologyor other methods to yield a random cross hatched pattern of lands 92 andgrooves 94. In this embodiment of the invention, the grooves form theair pockets or air reservoirs and the ratio of the groove area to theoverall bearing surface area should fall within the predetermined limitsrecited above.

Referring now to FIG. 17, FIG. 17A represents an enlarged, partiallycutaway perspective view of a portion of a steel shaft processed byconventional gauge pin manufacturing techniques to includecircumferential grooves 94 perpendicular to the axis of shaft rotation.FIG. 17B represents a partially cutaway sectional view of the lands 92and grooves 94 illustrated in FIG. 17A. In this application, althoughthe grooves are parallel to one another in a plane perpendicular to axisof rotation, the grooves are of random length and spacing. Such randomgroove distribution is an essential feature of the present invention.

In all of the embodiments described in FIGS. 15, 16 and 17, the groovesoccur randomly and create a near-infinite number of air pockets or airreservoirs within the overall surface area of the bearing assembly. Thisconfiguration of the invention is sharply distinct from the highlypatterned, repetitive and highly precise herringbone groove and landpattern described above in connection with FIG. 2.

To determine whether the necessary geometry limits, R_(a) limits, airreservoir ratio limits and aspect ratio limits have been met, a bearingassembly can be tested by placing the bearing assembly in an operatingapplication such as that illustrated in FIGS. 3 and 4 to determinewhether non-repeatable errors have been reduced to an acceptable levelfor specific applications. Non-repeatable errors of less than one arcsecond are routinely observed. These measurements can be madeelectronically or optically.

Each of the controllable parameters of the present invention isinherently related to the other parameters. For example, lowering theR_(a) rating of the bearing surfaces toward the eighteen R_(a) lowerlimit of the invention lowers the bearing stiffness and requiresimplementation of more accurate geometry tolerances to the sleeve/shaftgap to reduce the gap dimension toward the one hundred microinch lowerlimit. For higher bearing surface R_(a) ratings, much higher levels ofbearing stiffness are created enabling the use of looser bearing gapgeometric tolerances toward the upper limit of about three hundred.Similarly, as the ratio of air reservoir or pocket area to overallbearing surface area diminishes toward the lower limit, increased R_(a)ratings toward the upper limit can be used to compensate.

The numerous interrelationships between bearing gap geometry, R_(a)rating, air reservoir ratio and aspect ratio clearly demonstrate theempirical relationship between each of these parameters. Meeting theabove parameters are necessary to make the present invention work.

A process that is unnecessary to make the present invention work but maybe helpful to extend the start/stop cycle life is the application of adry lube which reduces the contact friction which occurs at the startand end of operation when the bearing is not airborne.

The random texture and closed flow design provides for low velocitypressurization and depressurization during the operating cycle.

For each of the different type of bearing surface components identifiedin Table 1, the method of implementing the required surface treatmentwill now be described in detail.

For ceramic sleeves or shafts, ceramics having an alumina content offrom about 94% to about 99.8% should be provided. Such materials areavailable from the Ceram Division of the Coors Ceramic Company of ElCahon, California or from the Mindrum Precision Products Company ofRancho Cucamonga, California. Either of these organizations can provideceramic surface finishes over a broad R_(a) range and can readilyprovide appropriate ceramic surface R_(a) ratings to specification.

The geometry of a ceramic shaft or sleeve must be controlled to acylindricity rating of twenty-five millionths of an inch. Thisspecification is defined by creating two concentric cylinders where thediameters of the inner cylinder is fifty millionths of an inch less thanthe diameters of the outer cylinder and where the gap between the outersurface of the inner cylinder and the inner surface of the outercylinder is equal to twenty-five millionths of an inch radially. Allparts of the surface of a ceramic element meeting this cylindricityspecification must fall within the gap between the two concentriccylinders. Ceramic bearing elements meeting this cylindricityspecification also meet all of the relevant geometric parameters ofstraightness, roundness and size uniformity.

The ratio of air pockets to overall bearing surface area can beinspected by microphotographs with an enlargement of between about300×to 600X. Optimum performance is achieved when this ratio equalsapproximately forty to fifty percent.

When all of the above-stated parameters have been achieved, no furthertreatment of ceramic bearing surfaces is necessary.

When either the bearing sleeve or shaft is fabricated from steel, fourhundred and forty (440) stainless steel or its equivalent has found tofunction acceptably. A cylindricity specification identical to thatdescribed above in connection with ceramic materials adequately controlsthe bearing surface geometry.

One acceptable method of manufacturing a steel sleeve capable offunctioning as a bearing element of the present invention involvesimplementation of the following sequence of steps:

1. Machining the sleeve bore to a slightly undersized diameter;

2. Initially honing the sleeve bore to increase the sleeve diameter tothe desired diameter and to achieve the desired surface geometry; and

3. Completing a final honing step to achieve the desired R_(a) figurewithin the limits of the invention.

To prepare a 440-steel shaft for use in a bearing of the presentinvention, the following procedures may be implemented to achieve thatpurpose:

1. Manufacturing the steel shaft to desired geometry specifications byconventional gauge pin manufacturing techniques including roll lappingthe surface of the steel shaft to an R_(a) of between about two to four;and

2. Rotating the steel shaft on a lathe to roughen the surface of theshaft with 180 grit wet or dry sandpaper where the sandpaper contactsthe steel shaft in a first pass with a first lateral direction ofmovement and in a second pass with a second lateral direction ofmovement to get the desired random cross hatched pattern of the typeillustrated in FIG. 16A and increase the R_(a) from two to four toapproximately eighteen to thirty.

Although the surface of the steel shaft can be finished and polisheddown to a much lower an R_(a) rating, the foregoing process indicatesthat abrasive techniques including application of wet or dry sandpaperare implemented to form a cross hatched pattern of grooves and lands toachieve an R_(a) rating of from eighteen to thirty, a far higher R_(a)rating than that which could have been achieved were smoothness thedesired object of the invention.

The following steps may be carried out to treat an aluminum shaft tosuccessfully function in a bearing of the present invention:

1. Honing the aluminum shaft to achieve a slightly undersized geometricdimension or lathe cutting or machining the aluminum shaft to thedesired geometry;

2. Hard anodizing the surface of the shaft;

3. Finish honing the anodized shaft to achieve finished geometricdimension; and

4. Honing the anodized aluminum surface with approximately two strokesof a Sunnen honing stone to achieve desired surface texture as describedbelow.

To provide appropriate surface treatment for an aluminum sleeve, thefollowing steps can be implemented:

1. Machining the aluminum bore to a slightly undersized dimension;

2. Hard anodizing the aluminum bore which results in a dimensional buildup of the bore;

3. Honing the anodized aluminum bore to the desired geometry; and

4. Honing the anodized aluminum bore with approximately two strokes of aSunnen honing stone to achieve the desired surface texture (SunnenProducts Company, St. Louis, Missouri For Step 3 honing, use SunnenStone No. K12-A55 (aluminum oxide, 220 grit, hardness of 5). For Step 4honing, use Sunnen Stone No. K12-A47 (aluminum oxide, 150 grit, hardnessof 7).

Although a limited number of material treatment procedures have beendescribed above to achieve operative surface texture, any one or more ofthe essentially equivalent surface finishing techniques listed belowcould be implemented using existing techniques to achieve the requiredR_(a) rating and air reservoir ratios:

1. Honing;

2. Etching;

3. Centerless grinding;

4. Ion implanting;

5. Shot peening;

6. Two-step machining/etching;

7. Burnishing;

8. EDM (electric discharge machine);

9. Plasma coating; and

10. Other equivalent techniques.

In one embodiment of the invention, the following dimensions were foundto yield highly acceptable bearing performance;

1. shaft diameter: 0.4060 inches

2. sleeve bore diameter: 0.40625 inches

3. bearing clearance (radial): 0.000125 inches nominal.

For the various material combinations listed in Table 1, variousadvantages and disadvantages have been observed or noted. The mostreliable combination represents implementation of a steel sleeve on aceramic shaft. The advantages of this combination are as follows:

1. Closely matched thermal coefficients of expansion;

2. A sleeve made from hardened 440C stainless steel can also function asthe rotor of a hysteresis synchronous motor, avoiding the requirementfor a permanent magnet rotor; and

3. High resistance to corrosion.

As to the combination of a hard anodized aluminum sleeve on a hardenedsteel shaft, the following advantages and disadvantages have beenobserved:

1. Steel shafts can be fabricated by well known and easily implementedgauge pin manufacturing techniques at reasonable cost to yield a surfacetextured shaft with excellent geometry;

2. The aluminum sleeve is easily machined and can be hard anodized toyield a hard, readily honed bearing surface; and

3. The dissimilar coefficients of thermal expansion of aluminum andsteel limit the operating temperature range of the bearing.

The following advantages and disadvantages have been noted fromobservation of bearings fabricated using hard anodized aluminum for boththe sleeve and shaft:

1. The identical thermal coefficient of expansion enables operation overwide temperature ranges; and

2. Aluminum materials easily machined.

As to the use of a ceramic sleeve and shaft, the following advantagesand disadvantages have been noted:

1. Identical thermal coefficient of expansion permits operation overwide temperature ranges; and

2. The material is a highly stable material.

Dry lubricant can by applied to one or both of the bearing surfaces tominimize frictional wear from commencement of bearing rotation untilbearing liftoff occurs at approximately seventy-five to two hundred andtwenty-five surface feet per minute. Between 0 velocity and liftoffvelocity during both start-up and shut down, the bearing surfacescontact each other and function as a contact bearing. Dry lubricantfunctions exclusively during this transition velocity region and reducesfrictional wear of the two contacting bearing surfaces. Dry lubricant isunnecessary to cause the present invention to function.

To limit contact phase operation bearing surface wear, the sleeve boreand shaft may be treated with tungsten disulfide or a boron nitritebased dry lubricant.

For a bearing including a ceramic sleeve and a ceramic shaft, drylubricant is typically not used. For the steel sleeve/ceramic shaftembodiments, dry lubricant may be applied to the sleeve bore. Foraluminum/steel bearing surfaces, lubricant may be applied to bothsurfaces.

The present invention can be implemented in various other embodiments inaddition to the specific embodiments described above. For example, FIG.18 illustrates the use of two counter-rotating sleeves on a common fixedshaft. A series of magnetic axial thrust bearings are coupled as shownto maintain the requisite axial alignment of the various bearingelements.

Referring now to FIG. 19, a spherical rotating bearing element 100 isrigidly coupled to motor shaft 102 and interfaces with a matching,stationary spherical bearing surface 104. In the FIG. 19 embodiment ofthe invention, an adjusting mechanism must be provided to adjust therelative position of rotating bearing surfaces 100 with respect tolongitudinal axis 106 to provide an appropriate gap dimension asexplained above. Such adjustment could be accomplished by providing anadjustable hub, adjustment screws or even by appropriately selectedshims.

Referring now to FIG. 20, the bearing assemblies of the presentinvention are fabricated as conic sections including a rotating conicbearing surface 108 and a stationary conic bearing surface 110. Bearingsurface 108 is rigidly coupled to a rotating shaft 112. As was the casewith the spheric section bearing assembly illustrated in FIG. 19,adjustment of at least one of the two bearing surfaces 108 with respectto longitudinal axis 112 must be provided by appropriate adjusting meanssuch as shims, screws or adjustable hubs.

In both the FIG. 19 and 20 embodiments of the invention, the rotatingand stationary bearing surfaces must be carefully matched as, forexample, by applying a fine grit abrasive material between the rotatingsurfaces to wear in and match the adjacent bearing surfaces. It may alsobe possible by appropriate, highly accurate machining techniques toavoid such an abrasive wearing step. Upon completion of such abrasivematching procedures, the resulting bearing surface texture must beevaluated to determine whether further bearing surface treatment isrequired to provide the appropriate R_(a) surface roughness required forappropriate operation of the present invention.

Referring now to FIG. 21A, the bearing assemblies of the presentinvention are fabricated as annular disc sections including rotatingdiscs 114 with bearing surface 116 and stationary discs 118 with bearingsurfaces 120. Rotating discs 114 are rigidly coupled to a rotatingsleeve 122. A cylindrical bearing shaft 124 is positioned coaxiallywithin bearing sleeve 122 having ends which are rigidly coupled togetherwith disc 118 to housing 126.

The bearing illustrated in FIG. 21A functions as an axial thrust controlbearing. The discs are matched to provide an appropriate gap dimensionand surface R_(a) as explained above.

A bearing of this configuration with thrust bearings on each end asshown in FIG. 21A can operate horizontally or vertically. A bearing withthrust surfaces on one end only can be operated in a vertical attitudewith the thrust surfaces supporting a load.

FIG. 21B illustrates a bearing that includes discs 128 with thrustsurfaces 130 rigidly coupled to rotating shaft 132. Disc 134 with thrustsurfaces 136 is rigidly coupled along with bearing sleeve 138 to thehousing 140.

FIG. 21C illustrates a bearing that includes disc 142 with thrustsurfaces 144 rigidly coupled to rotating sleeve 146. Discs 148 arerigidly coupled along with bearing shaft 150 to the housing 152.

Although only specific embodiments of these axial thrust controlbearings are illustrated in FIGS. 21A, B and C, the application of gassupported air bearings as axial thrust control bearing assemblies couldbe modified in numerous ways readily understandable to one of ordinaryskill in the art based upon principals illustrated in FIG. 21.

Numerous benefits are achieved by implementation of the presentinvention. The unique and interrelated combination of bearing geometry,R_(a) roughness, ratio of air reservoirs to bearing surface area andaspect ratio create an extraordinarily high bearing stiffness on theorder of 30,000 to 50,000 pounds per inch.

The unique structure of the present invention also results in a rapidstiffness build up as operating velocity increases from start up andresults in extremely low speed lift off of one bearing surface relativeto the other. The present invention experiences lift off at fromapproximately seventy-five to two hundred and twenty-five surface feetper minute. One prior art herringbone bearing assembly does not attainliftoff until approximately six hundred surface feet per minute.

The bearing of the present invention can also operate at extremely highRPM's. A prototype of the present invention was successfully tested at40,000 RPM, the maximum RPM of the prototype drive motor. Prior artherringbone bearings are typically limited to maximum RPM operation aton the order of about 30,000 RPM.

The extraordinarily high bearing stiffness ratings achieved by thepresent invention (on the order of 30,000 to 50,000 pounds per inch)permits operation of the bearing in any attitude including horizontal,vertical or inclined. The bearing stiffness ratings of some prior artherringbone systems are insufficient, require operation in a verticalattitude and cannot successfully operate for any significant amount oftime with inclinations of even ten degrees away from vertical.

The bearing of the present invention can also be operated with either aclockwise or counterclockwise rotation direction. The prior artherringbone bearings are unidirectional in view of the uniqueherringbone groove pattern and the requirement to pump air in a singledirection to pressurize the bearing.

The present invention operates as a closed system without a requirementfor an external air supply. Herringbone bearing assemblies require asource of air which is pumped through the bearing. Unless operated in adebris free environment, herringbone air bearings are nearly alwayscontaminated by airborne debris, causing catastrophic bearing failure.

The bearing of the present invention can operate at high altitude andhas been tested at altitudes up to 20,000 feet without significantperformance degradation.

The unique structure of the bearing of the present invention results inextremely low bearing surface wear. A prototype of the present inventionhas been tested for more than 40,000 start/stop cycles and althoughextremely high resolution measuring equipment was used to inspect forwear, no measurable wear could be discerned. The operating lifetime ofthis bearing is therefore predicted to be well in excess of 20,000start/stop cycles. Some prior art herringbone bearing assemblies aretypically specified as having a lifetime of only 10,000 start/stopcycles. A prototype of the present invention has been operatingcontinuously in excess of 22,000 hours to date with no evidence of wear.The operating lifetime of prior art ball bearing assemblies is typicallyon the order of about 2000 hours at speeds above 20,000 RPM.

The unique structure of the present invention provides a full lengthbearing assembly across the entire opposing surface of the bearingsleeve and the bearing shaft to provide very large bearing supportsurface areas providing excellent shock resistance to shipping andhandling damage. Prior art herringbone bearing assemblies rely onrelatively short, small area bearing surfaces.

If the materials used in the bearing sleeve and shaft are properlymatched with respect to thermal coefficients of expansion as can readilybe done, extremely wide temperature operating ranges can be readilyachieved.

Because the bearing assembly of the present invention experiencesextremely low frictional torque, a negligible heat rise is achieved,typically on the order of less than about 5° F. at an operating RPM ofabout 22,000 RPM. This limited heat rise is due primarily to motorheating.

The utilization of dry lubricant (when lubricant is used with thebearing of the present invention) instead of a wet lubricant as used inprior art ball bearing assemblies completely eliminates the lubricantcontamination problems experienced by prior art ball bearing units. Foroptical applications, lubricant contamination of the optical surfaces isthus completely avoided by use of the present invention.

The random redistribution of the wet lubricant used in prior art ballbearings causes random variations in the ball bearing drag forces. Theserandom redistributions contribute to rotational velocity error in therotating member. The elimination of the wet lubricant of the presentinvention gas bearing typically improves the velocity stability by afactor of 2 over prior art ball bearings.

Because the present invention utilizes a near-infinite number or randomlands and grooves and inexpensive manufacturing techniques to achievethe required bearing surface configuration, the bearing of the presentinvention can be manufactured at extremely low cost. The requirement ofprior art herringbone bearings for fixed, highly precise geometrypatterns result in close manufacturing tolerances and high manufacturingcosts.

It will be apparent to those skilled in the art that the disclosedself-pressurizing gas supported bearing may be modified in numerousother ways and may assume many other embodiments in addition to thepreferred forms specifically set out and described above.

During advanced research and development directed toward commercialapplication and mass production of the selfpressurizing gas supportedbearing with surface roughness finish as described above, it becameapparent that under specific operating conditions involving high bearingoperating RPM (25,000 RPM and higher) and low load (1/4 pound) thatalthough the invention exhibited superior performance in many ways incomparison to related prior art bearings, the bearing of the presentinvention frequently encountered an instability phenomenon commonlyreferred to as half-speed whirl instability. Such whirl instabilitymanifests itself as uncontrollable radial movement of the rotatingcylindrical bearing element relative to the stationary cylindricalbearing element, causing severe frictional erosion of the bearingsurfaces and rapid bearing failure. Under lower operating RPM conditionsor higher load conditions, the whirl instability problem described abovedid not arise.

Referring now to FIGS. 22-26, an improvement for eliminating this whirlinstability problem in either surface roughness bearings or conventionalsmooth bearings, will now be described in detail.

Referring initially to the embodiment of the invention illustrated inFIGS. 22, 23 and 26, a pneumatic load ramp 210 is formed as an eccentricannulus in the exterior surface of the midsection of a first cylindricalbearing element in the form of shaft 212. In this particular embodimentof the invention, a ceramic shaft 212 includes an overlap length 208 of2.68 inches while pneumatic load ramp 210 includes a length of 1.06inches. The term "overlap length" identifies the effective bearing areawhere the bearing surfaces of the shaft and sleeve overlap. Thisparticular configuration of the invention therefore divides bearingshaft 212 into undisturbed end surfaces 214 and 216 each having a lengthof 0.81 inches with each representing thirty percent of the overlaplength of shaft 212. In this embodiment, shaft 212 represents therotationally fixed or stationary bearing element. In an alternativeembodiment of the invention, shaft 212 could be dynamically balanced ina conventional manner and configured to serve as the rotating bearingelement.

Pneumatic load ramp 210 typically includes a circumferential sectiondesignated by reference number 218 where the surface of bearing shaft212 is not relieved or ground away. In one specific embodiment of theinvention, the width of unrelieved circumferential section 218 is equalto 0.30±0.03 inches measured around the circumference of shaft 212. Inthat same embodiment of the invention, the radius R₁ of shaft 212 equals0.2030±0.0005 inches while the radius R₂ of the eccentric annulus whichforms pneumatic load ramp 210 equals 0.2016±0.0005 inches. Thesedifferential radial dimensions typically yield a maximum relief of about0.003 to 0.004 inches between the circumference of pneumatic load ramp210 and the circumference of bearing shaft 212.

As illustrated in FIG. 26, the bearing of the present invention may beformed by positioning stationary shaft 212 coaxially within rotatingsleeve 220 where the clockwise direction of rotation of sleeve 220 isdesignated by reference number 222. The combined effect of the relativerotation between rotating sleeve 220 and stationary shaft 212 and thegeometric discontinuity created between the full radius section R₁ andthe reduced radius section R₂ of pneumatic load ramp 210 in combinationwith the internal airflow path designated by reference number 224creates an interaction in the form of an increased pressure air wedge226. Air wedge 226 maintains a fixed angular position immediatelycounterclockwise from unrelieved circumferential section 218 of shaft212. Air wedge 226 thereby creates a force or load across a relativelysmall included angle and along a less than full length section of thebearing surface. The angular position of the air wedge force remainsfixed relative to stationary bearing shaft 212.

This embodiment of the pneumatic load ramp of the present inventionthereby applies an asymmetric load to the bearing at a location fixedrelative to the stationary bearing shaft and provides a fixed load pointfor the bearing to react to. Since this load point does not and cannotmove relative to the stationary bearing shaft, the undesirablephenomenon of bearing whirl instability cannot arise as long as themagnitude of the asymmetric load generated by pneumatic load ramp 210 issufficient to resist the radially directed force vectors which otherwiseinduce bearing whirl instability.

Because the force generated by air wedge 226 increases with increasingbearing operating RPM and because the tendency of the bearing to createwhirl instability forces increases in a corresponding manner withincreasing bearing RPM, the force produced by air wedge 226, if of asufficient magnitude at a given bearing operating RPM, will typically beadequate to overcome bearing whirl instability forces regardless ofbearing operating RPM.

The magnitude of the radial force created by air wedge 226 is controlledby a variety of bearing geometry factors. The longitudinally orientedlength of pneumatic load ramp 218 is designated by reference number 228.The magnitude of the air wedge 226 force varies directly with the length228 of pneumatic load ramp 210. If the length of pneumatic load ramp 210is reduced below a critical minimum length, the bearing whirlinstability forces will overcome the counterbalancing force generated byair wedge 226 and the whirl instability problem will reappear. Thelength 228 of pneumatic load ramp must therefore be set at a minimumlength slightly in excess of this critical minimum length to ensuresatisfactory bearing operation.

As the length 228 of pneumatic load ramp 210 increases, the resultinglength of bearing end surfaces 214 and 216 decreases. As a criticalmaximum pneumatic load ramp length is exceeded, the total length ofbearing end surfaces 214 and 216 decreases below a critical minimumlength where air wedge 226 generates an asymmetric load in excess of theload which can be supported by bearing end surfaces 214 and 216. Thisexcessive pneumatic load ramp force asymmetrically displaces therotating bearing element relative to the stationary bearing element andresults in unsatisfactory bearing performance. Accordingly, forsatisfactory bearing operation, this critical maximum length of thepneumatic load ramp should not be exceeded.

Although the precise length 228 of pneumatic load ramp 210 is notparticularly critical, that length must be selected to lie between theminimum critical length and the maximum critical length described above.A highly satisfactory length for pneumatic load ramp 210 has been foundin the specific application described above to equal approximately onethird of the overall length of the bearing shaft. The length ofpneumatic load ramp 210 will typically occupy from between about 10% toabout 50% of the overlap bearing length 208 defined by the sum of thelengths of bearing end surfaces 214 and 216 and the length 228 ofpneumatic load ramp 210. Selection of a specific length 228 forpneumatic load ramp 210 in view of the teachings recited above regardingcritical minimum and maximum lengths would be readily apparent to one ofordinary skill in the bearing art.

Other factors relating to the geometry of the pneumatic load ramp of thepresent invention will also affect the magnitude of the force generatedby air wedge 226. For example, an appropriate length for the essentiallyunrelieved circumferential section 218 must be selected as well as thedifferential distance between radius R₁ of shaft 212 and radius R₂ ofthe pneumatic load ramp. Selection of these particular geometricparameters for specific bearing applications could be readily determinedby one of ordinary skill in the bearing art based on the teachingsrecited above.

Pneumatic load ramp 210 can readily be formed in the selected airbearing element by using conventional machine tools to grind the desiredgeometric configuration into the surface of the selected rotating orstationary bearing element. The geometric tolerances for the exteriorsurface dimensions of the pneumatic load ramp are not particularlycritical and can be empirically determined.

In an alternative embodiment of the invention illustrated in FIGS. 24and 25, the pneumatic load ramp of the present invention is formed as aneccentric annulus designated by reference number 230 recessed into thecylindrical interior surface of bearing sleeve 232. Determination of thepneumatic load ramp length and other geometric configurations such asthe circumferential length of the unrelieved circumferential section 234and radial distances R₁ and R₂ are determined as stated above inconnection with formation of the pneumatic load ramp in bearing shaft212. If the pneumatic load ramp 210 is formed in a bearing sleeve 232which serves as the rotating bearing element, the rotatingsleeve/pneumatic load ramp bearing element should be dynamicallybalanced using conventional techniques.

As illustrated in the drawings, the pneumatic load ramp of the presentinvention may be formed as either a reduced diameter eccentric sectionin the bearing shaft or as an increased diameter eccentric section inthe bearing sleeve.

The pneumatic load ramp of the present invention can be incorporated inself-pressurizing gas supported bearings including first and secondR_(a) roughness profiles falling within the range of from about eighteento about sixty as taught above. Alternatively, the pneumatic load rampof the present invention can be incorporated in conventional smooth orpolished air bearing surfaces having configurations and highly polishedsurface finish tolerances well known to those of ordinary skill in theart.

The inventive method of applying an asymmetric load to the bearing at alocation fixed relative to a selected stationary or rotating bearingsurface can be implemented with configurations different from thepneumatic load ramps illustrated in FIGS. 22-26. For example, anasymmetric load in compliance with the teachings of the presentinvention could be imparted to the bearing by positioning a single poleof a permanent magnet adjacent to a bearing having a rotatingferromagnetic steel sleeve and a stationary, non-ferromagnetic ceramicshaft. In another embodiment of the invention, a single pole of anelectromagnet could be positioned adjacent to a rotating ferromagneticsteel sleeve which coaxially surrounds a stationary, non-ferromagneticcylindrical shaft. In this alternative embodiment of the invention, theelectromagnet would be deenergized at bearing RPM's below which bearinglift-off occurs, such as during starting and stopping, to preventundesirable bearing surface wear caused by the asymmetric force inducedbetween the two bearing elements by the magnetic force on theferromagnetic bearing element. After bearing lift-off has been achieved,the electromagnet can be energized to apply the required level ofasymmetric force to the ferromagnetic bearing element. Since increasedasymmetric loads are required as bearing operating RPM increases, theoperating parameters of the electromagnet could be readily controlled toprovide increased magnetically-induced asymmetric force levels at higherbearing operating RPM and decreased asymmetric forces at lower bearingoperating RPM, as long as the asymmetric force is maintained above thecritical minimum level required to eliminate bearing whirl instability.

Numerous additional benefits of the invention are realized in additionto the ability to eliminate bearing whirl instability. For example, thepneumatic load ramp of the present invention will work regardless of therelative direction of rotation of the rotating bearing element relativeto the stationary bearing element. In the FIG. 22 embodiment of theinvention, the coaxial cylindrical sleeve can be rotated in either acounterclockwise or clockwise direction and the pneumatic load ramp ofthe present invention will operate satisfactorily. For the clockwisesleeve direction of rotation as shown in FIG. 26, air wedge 226 willform to the left of unrelieved circumferential section 218. For theopposite counterclockwise direction of sleeve rotation, air wedge 226will form on the right side of circumferential section 218. In otherwords, air wedge 226 always forms upstream of the airflow designated byreference number 224 created by relative rotation between the rotatingbearing element and the stationary bearing element.

Another very significant benefit of the invention is that it operates ina closed system using only air entrapped between the bearing surfacesand does not require a flow of air pumped from a source external to thebearing. Such external air circulation requirements common with priorart bearing systems are highly undesirable since the circulation of airfrom external sources virtually always transfers particulatecontaminants and debris into the critical, close tolerance bearingsurface interface, resulting in substantially increased bearing wearleading to shortened bearing life and unpredictable bearing failuremodes.

Because the configuration of the pneumatic load ramp of the presentinvention is not geometrically complex nor are the tolerances highlycritical, the pneumatic load ramp can be inexpensively added to existingbearing designs using relatively non-precision tooling and cuttingtechniques.

Because the asymmetric force created by the pneumatic load ramp of thepresent invention increases in proportion to increasing bearingoperating RPM, a single geometric configuration will function acceptablyat highly variable bearing operating speeds, including very high bearingoperating speeds.

Although the pneumatic load ramp of the present invention has beenspecifically illustrated in the drawings as taking the form of a singleeccentric annulus, numerous different geometric configurations couldreadily be implemented by persons of ordinary skill in the bearingdesign art to create an air wedge or other means for applying anasymmetric load to the bearing at a location fixed relative to aselected bearing element. For example, the pneumatic load ramp could beformed as two, three or more spaced apart elements in either bearingelement as shown in FIG. 27. In another alternative embodiment, thepneumatic load ramp could be formed with one ramp in one bearing elementand with two spaced apart ramps in the opposing bearing element.Accordingly, it is intended by the appended claims to cover all suchmodifications of the invention which fall within the true spirit andscope of the invention.

We claim:
 1. A gas-supported bearing assembly comprising:a. acylindrical bearing sleeve having a longitudinal axis and a cylindricalinner surface including a first bearing surface; b. a cylindricalbearing shaft positioned coaxially within the bearing sleeve, separatedfrom the bearing by a gaseous medium, and having a cylindrical outersurface including a second bearing surface, wherein a relativerotational velocity is established between the bearing sleeve and thebearing shaft to generate a bearing supporting force through the gaseousmedium along a bearing overlap zone where the first bearing surfaceoverlaps the second bearing surface but is separated from the secondbearing surface by the gaseous medium to form first and second spacedapart gas bearings, the bearing overlap zone having a defined lengthalong the longitudinal axis; and c. a pneumatic load ramp formed in oneof the bearing surfaces creating a chamber between the first and secondbearings and having a length along the longitudinal axis less than thelength of the bearing overlap zone for generating a load in the gaseousmedium within the camber to apply an asymmetric load through the gaseousmedium in the chamber to the bearing assembly at a location fixedrelative to either the sleeve or the shaft to eliminate bearing whileinstability, the first and second gas bearings restricting axial flow ofthe gaseous medium along the longitudinal axis into or out of thechamber.
 2. The gas-supported bearing assembly of claim 1 wherein thepneumatic load ramp is formed in the bearing shaft.
 3. The gas-supportedbearing assembly of claim 1 wherein the pneumatic load ramp is formed inthe sleeve.
 4. The gas-supported bearing assembly of claim 1 wherein thepneumatic load ramp is formed in both the bearing shaft and the sleeve.5. The gas-supported bearing assembly of claim 4 having a defined lengthalong the longitudinal axis wherein the pneumatic load ramp furtherincludes first, second and third spaced apart pneumatic load ramps. 6.The gas supported bearing assembly of claim 1 wherein the pneumatic loadramp is formed in the bearing shaft by a reduced diameter eccentricsection.
 7. The gas supported bearing assembly of claim 1 wherein thepneumatic load ramp is formed in the sleeve by an increased diametereccentric section.
 8. The gas supported bearing assembly of claim ofclaim 1 wherein the length of the pneumatic load ramp is approximatelyequal to 1/3 of the length of the bearing overlap zone.
 9. The gassupported bearing assembly of claim 1 wherein the pneumatic load rampincludes a longitudinally extending, circumferential discontinuity inone of the bearing surfaces.
 10. The gas supported bearing assembly ofclaim 9 wherein the circumferential discontinuity is recessed into thebearing surface.
 11. The gas supported bearing assembly of claim 10wherein none of the pneumatic load ramp protrudes beyond the firstbearing surface or beyond the second bearing surface.
 12. Thegas-supported bearing assembly of claim 1 wherein:a. the cylindricalbearing sleeve includes a first bearing surface with a random surfacetexture having a first R_(a) roughness profile; and b. the cylindricalbearing shaft includes a random surface texture having a second R_(a)roughness profile, wherein the sum of the first and second R_(a)roughness profiles falls within the range of from about eighteen toabout sixty.
 13. The gas-supported bearing assembly of claim 12 whereinthe pneumatic load ramp is formed in the bearing shaft.
 14. Thegas-supported bearing assembly of claim 12 wherein the pneumatic loadramp is formed in the bearing sleeve.